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Copyright © 2006, ASHRAE
CHAPTER 1
LIQUID OVERFEED SYSTEMS
Overfeed System Operation ....................................................... 1.1
Refrigerant Distribution............................................................. 1.2
Oil in System .............................................................................. 1.3
Circulating Rate ......................................................................... 1.3
Pump Selection and Installation ................................................ 1.4
Controls...................................................................................... 1.5
1.
Evaporator Design ..................................................................... 1.6
Refrigerant Charge..................................................................... 1.6
Start-Up and Operation ............................................................. 1.6
Line Sizing.................................................................................. 1.7
Low-Pressure Receiver 
Sizing ...................................................................................... 1.7
VERFEED systems force excess liquid, either mechanically orOby gas pressure, through organized-flow evaporators, separate
it from the vapor, and return it to the evaporators.
Terminology
Low-pressure receiver. Sometimes referred to as an accumula-
tor, this vessel acts as the separator for the mixture of vapor and liq-
uid returning from the evaporators. A constant refrigerant level is
usually maintained by conventional control devices.
Pumping unit. One or more mechanical pumps or gas-operated
liquid circulators are arranged to pump overfeed liquid to the evap-
orators. The pumping unit is located below the low-pressure re-
ceiver.
Wet returns. These are connections between the evaporator out-
lets and low-pressure receiver through which the mixture of vapor
and overfeed liquid is drawn.
Liquid feeds. These are connections between the pumping unit
outlet and evaporator inlets.
Flow control regulators. These devices regulate overfeed flow
into the evaporators. They may be needle valves, fixed orifices, cal-
ibrated manual regulating valves, or automatic valves designed to
provide a fixed liquid rate.
Advantages and Disadvantages
The main advantages of liquid overfeed systems are high system
efficiency and reduced operating expenses. These systems have
lower energy cost and fewer operating hours because
• The evaporator surface is used efficiently through good refriger-
ant distribution and completely wetted internal tube surfaces.
• The compressors are protected. Liquid slugs resulting from fluc-
tuating loads or malfunctioning controls are separated from suc-
tion gas in the low-pressure receiver.
• Low-suction superheats are achieved where suction lines between
the low-pressure receiver and the compressors are short. This
minimizes discharge temperature, preventing lubrication break-
down and minimizing condenser fouling.
• With simple controls, evaporators can be hot-gas defrosted with
little disturbance to the system.
• Refrigerant feed to evaporators is unaffected by fluctuating ambi-
ent and condensing conditions. Flow control regulators do not
need to be adjusted after initial setting because overfeed rates are
not generally critical.
• Flash gas resulting from refrigerant throttling losses is removed at
the low-pressure receiver before entering the evaporators. This
gas is drawn directly to the compressors and eliminated as a factor
in system low-side design. It does not contribute to increased
pressure drops in the evaporators or overfeed lines.
The preparation of this chapter is assigned to TC 10.1, Custom Engineered
Refrigeration Systems.
• Refrigerant level controls, level indicators, refrigerant pumps, and
oil drains are generally located in equipment rooms, which are
under operator surveillance or computer monitoring.
• Because of ideal entering suction gas conditions, compressors last
longer. There is less maintenance and fewer breakdowns. The oil
circulation rate to the evaporators is reduced as a result of the
low compressor discharge superheat and separation at the low-
pressure receiver (Scotland 1963).
• Automatic operation is convenient.
The following are possible disadvantages:
• In some cases, refrigerant charges are greater than those used in
other systems.
• Higher refrigerant flow rates to and from evaporators cause liquid
feed and wet return lines to be larger in diameter than high-
pressure liquid and suction lines for other systems.
• Piping insulation, which is costly, is generally required on all feed
and return lines to prevent condensation, frosting, or heat gain.
• Installed cost may be greater, particularly for small systems or
those with fewer than three evaporators.
• Operation of the pumping unit requires added expenses that are
offset by the increased efficiency of the overall system.
• Pumping units may require maintenance.
• Pumps sometimes have cavitation problems caused by low avail-
able net positive suction pressure.
Generally, the more evaporators used, the more favorable the ini-
tial costs for liquid overfeed compared to a gravity recirculated or
flooded system (Scotland 1970). Liquid overfeed systems compare
favorably with thermostatic valve feed systems for the same reason.
For small systems, the initial cost for liquid overfeed may be higher
than for direct expansion.
Ammonia Systems. Easy operation and lower maintenance are
attractive features for even small ammonia systems. However, for
ammonia systems operating below −18°C evaporating temperature,
some manufacturers do not supply direct-expansion evaporators be-
cause of unsatisfactory refrigerant distribution and control problems.
OVERFEED SYSTEM OPERATION
Mechanical Pump
Figure 1 shows a simplified pumped overfeed system in which a
constant liquid level is maintained in a low-pressure receiver. A
mechanical pump circulates liquid through the evaporator(s). The
two-phase return mixture is separated in the low-pressure receiver.
Vapor is directed to the compressor(s). Makeup refrigerant enters
the low-pressure receiver by means of a refrigerant metering device.
Figure 2 shows a horizontal low-pressure receiver with a mini-
mum pump pressure, two service valves in place, and a strainer on
the suction side of the pump. Valves from the low-pressure receiver
to the pump should be selected for minimal pressure drop. The
strainer protects hermetic pumps when oil is miscible with the
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1.2 2006 ASHRAE Handbook—Refrigeration (SI)
 
refrigerant. It should have a free area twice the transverse cross-
sectional area of the line in which it is installed. With ammonia, con-
sider using a suction strainer. Open-drive pumps do not require
strainers. If no strainer is used, a dirt leg should be used to reduce the
risk of solids getting into the pump.
Generally, minimum pump pressure should be at least double the
net positive suction pressure to avoid cavitation. Liquid velocity to
the pump should not exceed 0.9 m/s. Net positive suction pressure
and flow requirements vary with pump type and design; consult the
pump manufacturer for specific requirements. The pump should be
evaluated over the full range of operation at low and high flow. Cen-
trifugal pumps have a flat curve and have difficulty with systems in
which discharge pressure fluctuates.
Gas Pump
Figure 3 shows a basic gas-pumped liquid overfeed system, with
pumping power supplied by gas at condenser pressure. In this sys-
tem, a level control maintains the liquid level in the low-pressure
receiver. There are two pumper drums; one is filled by the low-pres-
sure receiver, and the other is drained as hot gas pushes liquid from
the pumper drum to the evaporator. Pumper drum B drains when hot
gas enters the drum through valve B. To function properly, the
pumper drums must be correctly vented so they can fill during the
fill cycle.Another common arrangement is shown in Figure 4. In this
system, high-pressure liquid is flashed into a controlled-pressure
receiver that maintains constant liquid pressure at the evaporator
inlets, resulting in continuous liquid feed at constant pressure. Flash
gas is drawn into the low-pressure receiver through a receiver pres-
sure regulator. Excess liquid drains into a liquid dump trap from
Fig. 1 Liquid Overfeed with Mechanical Pump
Fig. 1 Liquid Overfeed with Mechanical Pump
Fig. 2 Pump Circulation, Horizontal Separator
Fig. 2 Pump Circulation, Horizontal Separator
the low-pressure receiver. Check valves and a three-way equalizing
valve transfer liquid into the controlled-pressure receiver during the
dump cycle. Refined versions of this arrangement are used for mul-
tistage systems.
REFRIGERANT DISTRIBUTION
To prevent underfeeding and excessive overfeeding of refriger-
ants, metering devices regulate the liquid feed to each evaporator
and/or evaporator circuit. An automatic regulating device continu-
ously controls refrigerant feed to the design value. Other common
devices are hand expansion valves, calibrated regulating valves, ori-
fices, and distributors.
It is time-consuming to adjust hand expansion valves to achieve
ideal flow conditions. However, they have been used with some suc-
cess in many installations before more sophisticated controls were
Fig. 3 Double Pumper Drum System
Fig. 3 Double-Pumper-Drum System
Fig. 4 Constant-Pressure Liquid Overfeed System
Fig. 4 Constant-Pressure Liquid Overfeed System
Liquid Overfeed Systems 1.3
 
available. One factor to consider is that standard hand expansion
valves are designed to regulate flows caused by the relatively high
pressure differences between condensing and evaporating pressure.
In overfeed systems, large differences do not exist, so valves with
larger orifices may be needed to cope with the combination of
increased refrigerant quantity and relatively small pressure differ-
ences. Caution is necessary when using larger orifices because con-
trollability decreases as orifice size increases.
Calibrated, manually operated regulating valves reduce some of
the uncertainties involved in using conventional hand expansion
valves. To be effective, the valves should be adjusted to the manu-
facturer’s recommendations. Because refrigerant in the liquid feed
lines is above saturation pressure, the lines should not contain flash
gas. However, liquid flashing can occur if excessive heat gains by
the refrigerant and/or high pressure drops build up in feed lines.
Orifices should be carefully designed and selected; once in-
stalled, they cannot be adjusted. They are generally used only for
top- and horizontal-feed multicircuit evaporators. Foreign matter
and congealed oil globules can restrict flow; a minimum orifice of
2.5 mm is recommended. With ammonia, the circulation rate may
have to be increased beyond that needed for the minimum orifice
size because of the small liquid volume normally circulated. Pumps
and feed and return lines larger than minimum may be needed. This
does not apply to halocarbons because of the greater liquid volume
circulated as a result of fluid characteristics.
Conventional multiple-outlet distributors with capillary tubes of
the type usually paired with thermostatic expansion valves have
been used successfully in liquid overfeed systems. Capillary tubes
may be installed downstream of a distributor with oversized orifices
to achieve the required pressure reduction and efficient distribution.
Existing gravity-flooded evaporators with accumulators can be
connected to liquid overfeed systems. Changes may be needed only
for the feed to the accumulator, with suction lines from the accu-
mulator connected to the system wet return lines. An acceptable
arrangement is shown in Figure 5. Generally, gravity-flooded
evaporators have different circuiting arrangements from overfeed
evaporators. In many cases, the circulating rates developed by ther-
mosiphon action are greater than those used in conventional over-
feed systems.
Example 1. Find the orifice diameter of an ammonia overfeed system with
a refrigeration load per circuit of 4.47 kW and a circulating rate of 7.
Evaporating temperature is –35°C, pressure drop across the orifice is
55 kPa, and the coefficient of discharge for the orifice is 0.61. The cir-
culation per circuit is 33.3 mL/s.
Solution: Orifice diameter may be calculated as follows:
(1)
Fig. 5 Liquid Overfeed System Connected on Common Sys-
tem with Gravity-Flooded Evaporators
Fig. 5 Liquid Overfeed System Connected on Common 
System with Gravity-Flooded Evaporators
d
Q
Cd
------
⎝ ⎠
⎜ ⎟
⎛ ⎞0.5 ρ
p
---
⎝ ⎠
⎜ ⎟
⎛ ⎞0.25
=
where
d = orifice diameter, mm
Q = discharge through orifice, mL/s
p = pressure drop through orifice, Pa
ρ = density of fluid at −35°C
= 683.7 kg/m3
Cd = coefficient of discharge for orifice
Note: As noted in the text, use a 2.5 mm diameter orifice to avoid
clogging.
OIL IN SYSTEM
Despite reasonably efficient compressor discharge oil separators,
oil finds its way into the system low-pressure sides. In ammonia
overfeed systems, most of this oil can be drained from low-pressure
receivers with suitable oil drainage facilities. In low-temperature
systems, a separate valved and pressure-protected, noninsulated oil
drain pot can be placed in a warm space at the accumulator. The
oil/ammonia mixture flows into the pot, and the refrigerant evapo-
rates. This arrangement is shown in Figure 6. At subatmospheric
pressures, high-pressure vapor must be piped into the oil pot to force
oil out. Because of oil’s low solubility in liquid ammonia, thick oil
globules circulate with the liquid and can restrict flow through strain-
ers, orifices, and regulators. To maintain high efficiency, oil should
be removed from the system by regular draining.
Except at low temperatures, halocarbons are miscible with oil.
Therefore, positive oil return to the compressor must be ensured.
There are many methods, including oil stills using both electric heat
and heat exchange from high-pressure liquid or vapor. Some
arrangements are discussed in Chapter 2. At low temperatures, oil
skimmers must be used because oil migrates to the top of the low-
pressure receiver.
Build-up of excessive oil in evaporators must not be allowed
because it rapidly decreases efficiency. This is particularly critical in
evaporators with high heat transfer rates associated with low vol-
umes, such as flake ice makers, ice cream freezers, and scraped-
surface heat exchangers. Because refrigerant flow rate is high,
excessive oil can accumulate and rapidly reduce efficiency.
CIRCULATING RATE
In a liquid overfeed system, the circulating number or rate is
the mass ratio of liquid pumped to amount of vaporized liquid. The
amount of liquid vaporized is based on the latent heat for the refrig-
erant at the evaporator temperature. The overfeed rate is the ratio of
liquid to vapor returning to the low-pressure receiver. When vapor
leaves an evaporator at saturated vapor conditions with no excess
liquid, the circulating rate is 1 and the overfeed rate is 0. With a
d
33.3
0.61
----------
⎝ ⎠
⎜ ⎟
⎛ ⎞
0.5
683.7
55 1000×
-------------------------
⎝ ⎠
⎜ ⎟
⎛ ⎞
0.25
2.47 mm= =
Fig. 6 Oil Drain Pot Connected to Low-Pressure Receiver
Fig. 6 Oil Drain Pot Connected to Low-Pressure Receiver
1.4 2006 ASHRAE Handbook—Refrigeration (SI)
 
circulating rate of 4, the overfeed rate at full load is 3; at no load, it
is 4. Most systems are designed for steady flow. With few excep-
tions, load conditions may vary, causing fluctuating temperatures
outside and within the evaporator. Evaporator capacities vary con-
siderably; with constant refrigerant flow to the evaporator, the over-
feed rate fluctuates.
For each evaporator, there is an ideal circulating rate for every
loading condition that gives the minimum temperature difference
and best evaporator efficiency (Lorentzen 1968; Lorentzen and
Gronnerud 1967). With few exceptions, it is impossibleto predict
ideal circulating rates or to design a plant for automatic adjustment
of the rates to suit fluctuating loads. The optimum rate can vary with
heat load, pipe diameter, circuit length, and number of parallel cir-
cuits to achieve the best performance. High circulating rates can
cause excessively high pressure drops through evaporators and wet
return lines. Return line sizing (see the section on Line Sizing) can
affect the ideal rates. Many evaporator manufacturers specify rec-
ommended circulating rates for their equipment. Rates in Table 1
agree with these recommendations.
Table 1 Recommended Minimum Circulating Rate
Refrigerant Circulating Rate*
Ammonia (R-717)
Downfeed (large-diameter tubes) 6 to 7
Upfeed (small-diameter tubes) 2 to 4
R-22, upfeed 3
R-134a 2
*Circulating rate of 1 equals evaporating rate.
Because of distribution considerations, higher circulating rates
are common with top-feed evaporators. In multicircuit systems,
refrigerant distribution must be adjusted to provide the best possible
results. Incorrect distribution can cause excessive overfeed or star-
vation in some circuits. Manual or automatic regulating valves can
control flow for the optimum or design value.
Halocarbon densities are about twice that of ammonia. If halocar-
bons R-22, R-134a, and R-502 are circulated at the same rate as
ammonia, they require 6 to 8.3 times more energy for pumping to the
same height than the less-dense ammonia. Because pumping energy
must be added to the system load, halocarbon circulating rates are usu-
ally lower than those for ammonia. Ammonia has a relatively high
latent heat of vaporization, so for equal heat removal, much less
ammonia mass must be circulated compared to halocarbons.
Although halocarbons circulate at lower rates than ammonia, the
wetting process in the evaporators is still efficient because of the liq-
uid and vapor volume ratios. For example, at –40°C evaporating
temperature, with constant flow conditions in the wet return connec-
tions, similar ratios of liquid and vapor are experienced with a cir-
culating rate of 4 for ammonia and 2.5 for R-22, R-502, and R-134a.
With halocarbons, some additional wetting is also experienced
because of the solubility of the oil in these refrigerants.
When bottom feed is used for multicircuit coils, a minimum feed
rate per circuit is not necessary because orifices or other distribution
devices are not required. The circulating rate for top-feed and hori-
zontal feed coils may be determined by the minimum rates from the
orifices or other distributors in use.
Figure 7 provides a method for determining the liquid refrigerant
flow (Niederer 1964). The charts indicate the amount of refrigerant
Fig. 7 Charts for Determining Rate of Refrigerant Feed (No Flash Gas)
Fig. 7 Charts for Determining Rate of Refrigerant Feed (No Flash Gas)
Liquid Overfeed Systems 1.5
 
vaporized in a 1 kW system with circulated operation having no
flash gas in the liquid feed line. The value obtained from the chart
may be multiplied by the desired circulating rate and total refriger-
ation to determine total flow.
Pressure drop through flow control regulators is usually 10 to
50% of the available feed pressure. Pressure at the outlet of the flow
regulators must be higher than the vapor pressure at the low-
pressure receiver by an amount equal to the total pressure drop of
the two-phase mixture through the evaporator, any evaporator
pressure regulator, and wet return lines. Pressure loss could be
35 kPa in a typical system. When using recommended liquid feed
sizing practices, assuming a single-story building, the frictional
pressure drop from pump discharge to evaporators is about 70 kPa.
Therefore, a pump for 140 to 170 kPa should be satisfactory in this
case, depending on the lengths and sizes of feed lines, quantity and
types of fittings, and vertical lift involved.
PUMP SELECTION AND INSTALLATION
Types of Pumps
Mechanical pumps, gas pressure pumping systems, and injector
systems are available for liquid overfeed systems.
Types of mechanical pump drives include open, semihermetic,
magnetic clutch, and hermetic. Rotor arrangements include positive
rotary, centrifugal, and turbine vane. Positive rotary and gear pumps
are generally operated at slow speeds up to 900 rpm. Whatever type
of pump is used, take care to prevent flashing at the pump suction
and/or within the pump itself.
Centrifugal pumps are typically used for larger volumes,
whereas semihermetic pumps are best suited for halocarbons at or
below atmospheric refrigerant saturated pressure. Regenerative
turbines are used with relatively high pressure and large swings in
discharge pressure.
Open pumps are fitted with a wide variety of packing or seals.
For continuous duty, a mechanical seal with an oil reservoir or a liq-
uid refrigerant supply to cool, wash, and lubricate the seals is com-
monly used. Experience with the particular application or the
recommendations of an experienced pump supplier are the best
guide for selecting the packing or seal. The motor and pump can be
magnetically coupled instead of shaft coupled to eliminate shaft
seals. A small immersion electric heater in the oil reservoir can be
used with low-temperature systems to ensure that the oil remains
fluid. Motors should have a service factor that compensates for drag
on the pump if the oil is cold or stiff.
Considerations should include ambient temperatures, heat leak-
age, fluctuating system pressures from compressor cycling, inter-
nal bypass of liquid to pump suction, friction heat, motor heat
conduction, dynamic conditions, cycling of automatic evaporator
liquid and suction stop valves, action of regulators, gas entrance
with liquid, and loss of subcooling by pressure drop. Another factor
to consider is the time lag caused by the heat capacity of pump suc-
tion, cavitation, and net positive suction pressure factors (Lorent-
zen 1963).
The motor and stator of hermetic pumps are separated from the
refrigerant by a thin nonmagnetic membrane. The metal membrane
should be strong enough to withstand system design pressures. Nor-
mally, the motors are cooled and the bearings lubricated by liquid
refrigerant bypassed from the pump discharge. It is good practice to
use two pumps, one operating and one standby.
Installing and Connecting Mechanical Pumps
Because of the sensitive suction conditions of mechanical pumps
on overfeed systems, the manufacturer’s application and installation
specifications must be followed closely. Suction connections should
be as short as possible, without restrictions, valves, or elbows.
Angle or full-flow ball valves should be used. Using valves with
horizontal valve spindles eliminates possible traps. Gas binding is
more likely with high evaporating pressures.
Installing discharge check valves prevents backflow. Relief valves
should be used, particularly for positive-displacement pumps. Strain-
ers are not usually installed in ammonia pump suction lines because
they plug with oil. Strainers, although a poor substitute for a clean
installation, protect halocarbon pumps from damage by dirt or pipe
scale.
Pump suction connections to liquid legs (vertical drop legs from
low-pressure receivers) should be made above the bottom of the legs
to allow collection space for solids and sludge. Consider using vor-
tex eliminators, particularly when submersion of the suction inlet is
insufficient to prevent the intake of gas bubbles. Lorentzen (1963,
1965) gives more complete information.
Sizing the pump suction line is important. The general velocity
should be about 0.9 m/s. Small lines cause restrictions; oversized
lines can cause bubble formation during evaporator temperature
decrease because of the heat capacity of the liquid and piping. Over-
sized lines also increase heat gain from ambient spaces. Oil heaters
for the seal lubrication system keep the oil fluid, particularly during
operation below −18°C. Thermally insulating all cold surfaces of
pumps, lines, and receivers increases efficiency.CONTROLS
The liquid level in the low-pressure receiver can be controlled by
conventional devices such as low-pressure float valves, combina-
tions of float switch and solenoid valve with manual regulator, ther-
mostatic level controls, electronic level sensors, or other proven
automatic devices. High-level float switches are useful in stopping
compressors and/or operating alarms; they are mandatory in some
areas. Solenoid valves should be installed on liquid lines (minimum
sized) feeding low-pressure receivers so that positive shutoff is
automatically achieved with system shutdown. This prevents exces-
sive refrigerant from collecting in low-pressure receivers, which can
cause carryover at start-up.
To prevent pumps from operating without liquid, low-level float
switches can be fitted on liquid legs. An alternative device, a differ-
ential pressure switch connected across pump discharge and suction
connections, stops the pump without interrupting liquid flow. Cav-
itation can also cause this control to operate. When hand expansion
valves are used to control the circulation rate to evaporators, the ori-
fice should be sized for operation between system high and low
pressures. Occasionally, with reduced inlet pressure, these valves
can starve the circuit. Calibrated, manually adjusted regulators are
available to meter the flow according to the design conditions. An
automatic flow-regulating valve specifically for overfeed systems is
available.
Liquid and suction solenoid valves must be selected for refriger-
ant flow rates by mass or volume, not by refrigeration ratings from
capacity tables. Evaporator pressure regulators should be sized
according to the manufacturer’s ratings for overfeed systems.
Notify the manufacturer that valves being ordered are for overfeed
application, because slight modifications may be required. When
evaporator pressure regulators are used on overfeed systems for
controlling air defrosting of cooling units (particularly when fed
with very-low-temperature liquid), refrigerant heat gain may be
achieved by sensible, not latent, effect. In such cases, other defrost-
ing methods should be investigated. The possibility of connecting
the units directly to high-pressure liquid should be considered, espe-
cially if the loads are minor.
When a check valve and a solenoid valve are paired on an overfeed
system liquid line, the check valve should be downstream from the
solenoid valve. When the solenoid valve is closed, dangerous hydrau-
lic pressure can build up from expansion of the trapped liquid as it is
heated. When evaporator pressure regulators are used, entering liquid
pressure should be high enough to cause flow into the evaporator.
1.6 2006 ASHRAE Handbook—Refrigeration (SI)
 
Multicircuit systems must have a bypass relief valve in the pump
discharge. The relief valve’s pressure should be set considering the
back pressure on the valve from the low-pressure receiver. For
example, if the low-pressure receiver is set at 300 kPa and the max-
imum discharge pressure from the pump is 900 kPa, the relief valve
should be set at 600 kPa. When some circuits are closed, excess liq-
uid is bypassed into the low-pressure receiver rather than forced
through the evaporators still in operation. This prevents higher evap-
orating temperatures from pressurizing evaporators and reducing
capacities of operating units. Where low-temperature liquid feeds
can be isolated manually or automatically, relief valves can be
installed to prevent damage from excessive hydraulic pressure.
EVAPORATOR DESIGN
Considerations
There is an ideal refrigerant feed and flow system for each evap-
orator design and arrangement. An evaporator designed for gravity-
flooded operation cannot always be converted to an overfeed
arrangement, and vice versa; neither can systems always be designed
to circulate the optimum flow rate. When top feed is used to ensure
good distribution, a minimum quantity per circuit must be circulated,
generally about 30 mL/s. In bottom-feed evaporators, distribution is
less critical than in top or horizontal feed because each circuit fills
with liquid to equal the pressure loss in other parallel circuits.
Circuit length in evaporators is determined by allowable pressure
drop, load per circuit, tubing diameter, overfeed rate, type of refrig-
erant, and heat transfer coefficients. The most efficient circuiting is
determined in most cases through laboratory tests conducted by the
evaporator manufacturers. Their recommendations should be fol-
lowed when designing systems.
Top Feed Versus Bottom Feed
System design must determine whether evaporators are to be top
fed or bottom fed, although both feed types can be installed in a sin-
gle system. Each feed type has advantages; no arrangement is best
for all systems.
Advantages of top feed include
• Smaller refrigerant charge
• Possibly smaller low-pressure receiver
• Possible absence of static pressure penalty
• Better oil return
• Quicker, simpler defrost arrangements
For halocarbon systems with greater fluid densities, the refriger-
ant charge, oil return, and static pressure are very important.
Bottom feed is advantageous in that
• Distribution considerations are less critical
• Relative locations of evaporators and low-pressure receivers are
less important
• System design and layout are simpler
The top-feed system is limited by the relative location of compo-
nents. Because this system sometimes requires more refrigerant cir-
culation than bottom-feed systems, it has greater pumping load,
possibly larger feed and return lines, and increased line pressure
drop penalties. In bottom-feed evaporators, multiple headers with
individual inlets and outlets can be installed to reduce static pressure
penalties. For high lift of return overfeed lines from the evaporators,
dual suction risers eliminate static pressure penalties (Miller 1974,
1979).
Distribution must be considered when using a vertical refrigerant
feed, because of static pressure variations in the feed and return header
circuits. For example, for equal circuit loadings in a horizontal-
airflow unit cooler, using gradually smaller orifices for bottom-feed
circuits than for upper circuits can compensate for pressure differ-
ences.
When the top-feed free-draining arrangement is used for air-
cooling units, liquid solenoid control valves can be used during the
defrost cycle. This applies in particular to air, water, or electric
defrost units. Any liquid remaining in the coils rapidly evaporates or
drains to the low-pressure receiver. Defrost is faster than in bottom-
feed evaporators.
REFRIGERANT CHARGE
Overfeed systems need more refrigerant than dry expansion sys-
tems. Top-feed arrangements have smaller charges than bottom-
feed systems. The amount of charge depends on evaporator volume,
circulating rate, sizes of flow and return lines, operating tempera-
ture differences, and heat transfer coefficients. Generally, top-feed
evaporators operate with the refrigerant charge occupying about 25
to 40% of the evaporator volume. The refrigerant charge for the bot-
tom-feed arrangement occupies about 60 to 75% of the evaporator
volume, with corresponding variations in the wet returns. Under
some no-load conditions in up-feed evaporators, the charge may
occupy 100% of the evaporator volume. In this case, the liquid surge
volume from full load to no load must be considered in sizing the
low-pressure receiver (Miller 1971, 1974).
Evaporators with high heat transfer rates, such as flake ice mak-
ers and scraped-surface heat exchangers, have small charges
because of small evaporator volumes. The amount of refrigerant in
the low side has a major effect on the size of the low-pressure
receiver, especially in horizontal vessels. The cross-sectional area
for vapor flow in horizontal vessels is reduced with increasing liquid
level. It is important to ascertain the evaporator refrigerant charge
with fluctuating loads for correct vessel design, particularly for a
low-pressure receiver thatdoes not have a constant level control but
is fed through a high-pressure control.
START-UP AND OPERATION
All control devices should be checked before start-up. If mechan-
ical pumps are used, the direction of operation must be correct. Sys-
tem evacuation and charging procedures are similar to those for other
systems. The system must be operating under normal conditions to
determine the total required refrigerant charge. Liquid height is
established by liquid level indicators in the low-pressure receivers.
Calibrated, manually operated regulators should be set for the
design conditions and adjusted for better performance when neces-
sary. When hand expansion valves are used, the system should be
started by opening the valves about one-quarter to one-half turn.
When balancing is necessary, the regulators should be cut back on
circuits not starved of liquid, to force the liquid through underfed
circuits. The outlet temperature of the return line from each evapo-
rator should be the same as the main return line’s saturation temper-
ature, allowing for pressure drops. Starved circuits are indicated by
temperatures higher than those for adequately fed circuits. Exces-
sive feed to a circuit increases evaporator temperature because of
excessive pressure drop.
The relief bypass from the liquid line to the low-pressure receiver
should be adjusted and checked to ensure that it is functioning. Dur-
ing operation, the pump manufacturer’s recommendations for lubri-
cation and maintenance should be followed. Regular oil draining
procedures should be established for ammonia systems; the quanti-
ties of oil added to and drained from each system should be com-
pared, to determine whether oil is accumulating. Oil should not be
drained in halocarbon systems. Because of oil’s miscibility with
halocarbons at high temperatures, it may be necessary to add oil to
the system until an operating balance is achieved (Soling 1971;
Stoecker 1960).
Operating Costs and Efficiency
Operating costs for overfeed systems are generally lower than for
other systems (though not always, because of the various inefficien-
Liquid Overfeed Systems 1.7
 
LOW-PRESSURE RECEIVERSIZINGLOW-PRESSURE RECEIVER SIZING
cies that exist from system to system and from plant to plant). For
existing dry expansion plants converted to liquid overfeed, the op-
erating hours, power, and maintenance costs are reduced. Efficiency
of the early gas pump systems has been improved by using high-side
pressure to circulate overfeed liquid. This type of system is indi-
cated in the controlled-pressure system shown in Figure 4. Refine-
ments of the double-pumper-drum arrangement (see Figure 3) have
also been developed.
Gas-pumped systems, which use refrigerant gas to pump liquid
to the evaporators or to the controlled-pressure receiver, require
additional compressor volume, from which no useful refrigeration
is obtained. These systems consume 4 to 10% or more of the com-
pressor power to maintain refrigerant flow.
If condensing pressure is reduced as much as 70 kPa, the com-
pressor power per unit of refrigeration drops by about 7%. Where
outdoor dry- and wet-bulb conditions allow, a mechanical pump can
be used to pump gas with no effect on evaporator performance. Gas-
operated systems must, however, maintain the condensing pressure
within a much smaller range to pump the liquid and maintain the
required overfeed rate.
LINE SIZING
The liquid feed line to the evaporator and wet return line to the
low-pressure receiver cannot be sized by the method described in
Chapter 36 of the 2005 ASHRAE Handbook—Fundamentals. Fig-
ure 7 can be used to size liquid feed lines. The circulating rate from
Table 1 is multiplied by the evaporating rate. For example, an evap-
orator with a circulating rate of 4 that forms vapor at a rate of 50 g/s
needs a feed line sized for 4 × 50 = 200 g/s.
Alternative methods that may be used to design wet returns
include the following:
• Use one pipe size larger than calculated for vapor flow alone.
• Use a velocity selected for dry expansion reduced by the factor
. This method suggests that the wet-return
velocity for a circulating rate of 4 is = 0.5, or half that of the
acceptable dry-vapor velocity.
• Use the design method described by Chaddock et al. (1972). The
report includes tables of flow capacities at 0.036 K drop per metre
of horizontal lines for R-717 (ammonia), R-12, R-22, and R-502.
When sizing refrigerant lines, the following design precautions
should be taken:
• Carefully size overfeed return lines with vertical risers because
more liquid is held in risers than in horizontal pipe. This holdup
increases with reduced vapor flow and increases pressure loss
because of gravity and two-phase pressure drop.
• Use double risers with halocarbons to maintain velocity at partial
loads and to reduce liquid static pressure loss (Miller 1979).
• Add the equivalent of a 100% liquid static height penalty to the
pressure drop allowance to compensate for liquid holdup in
ammonia systems that have unavoidable vertical risers.
• As alternatives in severe cases, provide traps and a means of
pumping liquids, or use dual-pipe risers.
• Install low-pressure drop valves so the stems are horizontal or
nearly so (Chisholm 1971).
Low-pressure receivers are also called liquid separators, suction
traps, accumulators, liquid/vapor separators, flash coolers, gas and
liquid coolers, surge drums, knockout drums, slop tanks, or low-side
pressure vessels, depending on their function and user preference.
Low-pressure receiver sizing is determined by the required liquid
holdup volume and allowable gas velocity. The volume must
accommodate fluctuations of liquid in the evaporators and overfeed
return lines as a result of load changes and defrost periods. It must
1/Circulating Rate
1/4
also handle swelling and foaming of the liquid charge in the
receiver, which is caused by boiling during temperature increase or
pressure reduction. At the same time, a liquid seal must be main-
tained on the supply line for continuous circulation devices. A sep-
arating space must be provided for gas velocity low enough to cause
a minimum entrainment of liquid drops into the suction outlet.
Space limitations and design requirements result in a wide variety of
configurations (Lorentzen 1966; Miller 1971; Niemeyer 1961;
Scheiman 1963, 1964; Sonders and Brown 1934; Stoecker 1960;
Younger 1955).
In selecting a gas-and-liquid separator, adequate volume for the
liquid supply and a vapor space above the minimum liquid height
for liquid surge must be provided. This requires analysis of operat-
ing load variations. This, in turn, determines the maximum oper-
ating liquid level. Figures 8 and 9 identify these levels and the
important parameters of vertical and horizontal gravity separators.
Vertical separators maintain the same separating area with level
variations, whereas separating areas in horizontal separators change
with level variations. Horizontal separators should have inlets and
outlets separated horizontally by at least the vertical separating
distance. A useful arrangement in horizontal separators distributes
the inlet flow into two or more connections to reduce turbulence and
horizontal velocity without reducing the residence time of the gas
flow within the shell (Miller 1971).
In horizontal separators, as the horizontal separating distance
increases beyond the vertical separating distance, the residence time
of vapor passing through increases so that higher velocities than
allowed in vertical separators can be tolerated. As the separating
distance reduces, the amount of liquid entrainment from gravity
separators increases. Table 2 shows the gravity separation veloci-
ties. For surging loads or pulsating flow associated with large step
changes in capacity, the maximum steady-flow velocity should be
reduced to a value achieved by a suitable multiplier such as 0.75.
The gas-and-liquid separator may be designed with baffles or
eliminators to separate liquid fromthe suction gas returning from
the top of the shell to the compressor. More often, enough separation
space is allowed above the liquid level for this purpose. Such a
design is usually of the vertical type, with a separation height above
the liquid level of 600 to 900 mm. The shell diameter is sized to keep
suction gas velocity low enough to allow liquid droplets to separate
and not be entrained with the returning suction gas off the top of the
shell.
Although separators are made with length-to-diameter (L/D)
ratios of 1/1 increasing to 10/1, the least expensive separators usu-
ally have L /D ratios between 3/1 and 5/1. Vertical separators are
normally used for systems with reciprocating compressors. Hori-
zontal separators may be preferable where vertical height is critical
and/or where large volume space for liquid is required. The proce-
dures for designing vertical and horizontal separators are different.
A vertical gas-and-liquid separator is shown in Figure 9. The end
of the inlet pipe C1 is capped so that flow dispersion is directed
down toward the liquid level. The suggested opening is four times
the transverse internal area of the pipe. Height H1 with a 120° dis-
Fig. 8 Basic Horizontal Gas-and-Liquid Separator
Fig. 8 Basic Horizontal Gas-and-Liquid Separator
1.8 2006 ASHRAE Handbook—Refrigeration (SI)
 
persion of the flow reaches approximately 70% of the internal diam-
eter of the shell.
An alternative inlet pipe with a downturned elbow or mitered
bend can be used. However, the jet effect of entering fluid must be
considered to avoid undue splashing. The pipe outlet must be a min-
imum distance of IDS/5 above the maximum liquid level in the
Table 2 Maximum Effective Separation Velocities for R-717, 
R-22, R-12, and R-502, with Steady Flow Conditions
Temp., 
°C
Vertical
Separation
Distance, mm
Maximum Steady Flow Velocity, m/s
R-717 R-22 R-12 R-502
+10 250 0.15 0.07 0.08 0.06
610 0.64 0.31 0.36 0.25
910 0.71 0.39 0.43 0.32
−7 250 0.21 0.10 0.11 0.08
610 0.87 0.44 0.49 0.35
910 0.99 0.52 0.58 0.42
−23 250 0.31 0.14 0.16 0.11
610 1.29 0.61 0.69 0.49
910 1.43 0.72 0.81 0.59
−40 250 0.48 0.21 0.24 0.17
610 1.99 0.88 1.01 0.71
910 2.17 1.04 1.17 0.84
−57 250 0.80 0.33 0.37 0.25
610 3.30 1.36 1.54 1.08
910 3.54 1.57 1.78 1.25
Source: Adapted from Miller (1971).
Fig. 9 Basic Vertical Gravity Gas and
Liquid Separator
Fig. 9 Basic Vertical Gravity Gas and Liquid Separator
shell. H2 is measured from the outlet to the inside top of the shell. It
equals D + 0.5 times the depth of the curved portion of the head.
For the alternative location of C2, determine IDS from the fol-
lowing equation:
(2)
The maximum liquid height in the separator is a function of the
type of system in which the separator is being used. In some systems
this can be estimated, but in others, previous experience is the only
guide for selecting the proper liquid height. Accumulated liquid
must be returned to the system by a suitable means at a rate compa-
rable to its collection rate.
With a horizontal separator, the vertical separation distance used
is an average value. The top part of the horizontal shell restricts gas
flow so that the maximum vertical separation distance cannot be
used. If Ht represents the maximum vertical distance from the liquid
level to the inside top of the shell, the average separation distance as
a fraction of IDS is as follows:
The suction connection(s) for refrigerant gas leaving the hori-
zontal shell must be at or above the location established by the aver-
age distance for separation. The maximum cross-flow velocity of
gas establishes residence time for the gas and any entrained liquid
droplets in the shell. The most effective removal of entrainment
occurs when residence time is the maximum practical. Regardless
of the number of gas outlet connections for uniform distribution of
gas flow, the cross-sectional area of the gas space is
(3)
where
Ax = minimum transverse net cross-sectional area or gas space, mm
2
D = average vertical separation distance, mm
Q = total quantity of gas leaving vessel, L/s
L = inside length of shell, mm
V = separation velocity for separation distance used, m/s
For nonuniform distribution of gas flow in the horizontal shell,
determine the minimum horizontal distance for gas flow from point
of entry to point of exit as follows:
(4)
where
RTL = residence time length, mm
Q = maximum flow for that portion of the shell, L/s
All connections must be sized for the flow rates and pressure drops
permissible and must be positioned to minimize liquid splashing.
Internal baffles or mist eliminators can reduce vessel diameter;
however, test correlations are necessary for a given configuration
and placement of these devices.
An alternative formula for determining separation velocities that
can be applied to separators is
(5)
Ht /IDS D/IDS Ht /IDS D/IDS
0.1 0.068 0.6 0.492
0.2 0.140 0.7 0.592
0.3 0.215 0.8 0.693
0.4 0.298 0.9 0.793
0.5 0.392 1.0 0.893
IDS 1270Q
V
---------------- C2
2
+=
Ax
2000DQ
VL
---------------------=
RTL
1000QD
VAx
---------------------=
v k
ρ l ρv–
ρv
-----------------=
Liquid Overfeed Systems 1.9
 
where
v = velocity of vapor, m/s
ρl = density of liquid, kg/m3
ρv = density of vapor, kg/m3
k = factor based on experience without regard to vertical separation 
distance and surface tension for gravity separators
In gravity liquid/vapor separators that must separate heavy
entrainment from vapors, use a k of 0.03. This gives velocities
equivalent to those used for 300 to 350 mm vertical separation dis-
tance for R-717 and 350 to 400 mm vertical separation distance for
halocarbons. In knockout drums that separate light entrainment, use
a k of 0.06. This gives velocities equivalent to those used for
900 mm vertical separation distance for R-717 and for halocarbons.
REFERENCES
Chaddock, J.B., D.P. Werner, and C.G. Papachristou. 1972. Pressure drop in
the suction lines of refrigerant circulation systems. ASHRAE Trans-
actions 78(2):114-123.
Chisholm, D. 1971. Prediction of pressure drop at pipe fittings during two-
phase flow. Proceedings of the IIR Conference, Washington, D.C.
Lorentzen, G. 1963. Conditions of cavitation in liquid pumps for refrigerant
circulation. Progress Refrigeration Science Technology I:497.
Lorentzen, G. 1965. How to design piping for liquid recirculation. Heating,
Piping & Air Conditioning (June):139.
Lorentzen, G. 1966. On the dimensioning of liquid separators for refrigera-
tion systems. Kältetechnik 18:89.
Lorentzen, G. 1968. Evaporator design and liquid feed regulation. Journal
of Refrigeration (November-December):160.
Lorentzen, G. and R. Gronnerud. 1967. On the design of recirculation type
evaporators. Kulde 21(4):55.
Miller, D.K. 1971. Recent methods for sizing liquid overfeed piping and
suction accumulator-receivers. Proceedings of the IIR Conference,
Washington, D.C.
Miller, D.K. 1974. Refrigeration problems of a VCM carrying tanker. ASH-
RAE Journal 11.
Miller, D.K. 1979. Sizing dual suction risers in liquid overfeed refrigeration
systems. Chemical Engineering 9.
Niederer, D.H. 1964. Liquid recirculation systems—What rate of feed is rec-
ommended. The Air Conditioning & Refrigeration Business (December).
Niemeyer, E.R. 1961. Check these points when designing knockout drums.
Hydrocarbon Processing and Petroleum Refiner (June).
Scheiman, A.D. 1963. Size vapor-liquid separators quicker by nomograph.
Hydrocarbon Processing and Petroleum Refiner (October).
Scheiman, A.D. 1964. Horizontal vapor-liquid separators. Hydrocarbon
Processing and Petroleum Refiner (May).
Scotland, W.B. 1963. Discharge temperature considerations with multi-
cylinder ammonia compressors. Modern Refrigeration (February).
Scotland, W.B. 1970. Advantages, disadvantages and economics of liquid
overfeed systems. ASHRAE Symposium Bulletin KC-70-3: Liquid over-
feed systems.
Soling, S.P. 1971. Oil recovery from low temperature pump recirculating
hydrocarbon systems. ASHRAE Symposium Bulletin PH-71-2: Effectof oil on the refrigeration system.
Sonders, M. and G.G. Brown. 1934. Design of fractionating columns, en-
trainment and capacity. Industrial & Engineering Chemistry (January).
Stoecker, W.F. 1960. How to design and operate flooded evaporators for
cooling air and liquids. Heating, Piping & Air Conditioning (December).
Younger, A.H. 1955. How to size future process vessels. Chemical Engi-
neering (May).
BIBLIOGRAPHY
Chaddock, J.B. 1976. Two-phase pressure drop in refrigerant liquid overfeed
systems—Design tables. ASHRAE Transactions 82(2):107-133.
Chaddock, J.B., H. Lau, and E. Skuchas. 1976. Two-phase pressure drop in
refrigerant liquid overfeed systems—Experimental measurements. ASH-
RAE Transactions 82(2):134-150.
Geltz, R.W. 1967. Pump overfeed evaporator refrigeration systems. Air Con-
ditioning, Heating & Refrigeration News (January 30, February 6, March
6, March 13, March 20, March 27).
Lorentzen, G. and A.O. Baglo. 1969. An investigation of a gas pump recir-
culation system. Proceedings of the Xth International Congress of
Refrigeration, p. 215. International Institute of Refrigeration, Paris.
Richards, W.V. 1959. Liquid ammonia recirculation systems. Industrial
Refrigeration (June):139.
Richards, W.V. 1970. Pumps and piping in liquid overfeed systems.
ASHRAE Symposium Bulletin KC-70-3: Liquid overfeed systems.
Slipcevic, B. 1964. The calculation of the refrigerant charge in refrigerating
systems with circulation pumps. Kältetechnik 4:111.
Thompson, R.B. 1970. Control of evaporators in liquid overfeed systems.
ASHRAE Symposium Bulletin KC-70-3: Liquid overfeed systems.
Watkins, J.E. 1956. Improving refrigeration systems by applying established
principles. Industrial Refrigeration (June).
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	Overfeed System Operation 
	Refrigerant Distribution 
	Oil in System 
	Circulating Rate 
	Pump Selection and Installation 
	Controls 
	Evaporator Design 
	Refrigerant Charge 
	Start-Up and Operation 
	Line Sizing 
	Low-Pressure Receiver Sizing 
	References 
	Bibliography 
	Tables
	 1 Recommended Minimum Circulating Rate 
	 2 Maximum Effective Separation Velocities for R-717, R-22, R-12, and R-502, with Steady Flow Conditions 
	Figures
	 1 Liquid Overfeed with Mechanical Pump 
	 2 Pump Circulation, Horizontal Separator 
	 3 Double Pumper Drum System 
	 4 Constant-Pressure Liquid Overfeed System 
	 5 Liquid Overfeed System Connected on Common System with Gravity-Flooded Evaporators 
	 6 Oil Drain Pot Connected to Low-Pressure Receiver 
	 7 Charts for Determining Rate of Refrigerant Feed (No Flash Gas) 
	 8 Basic Horizontal Gas-and-Liquid Separator 
	 9 Basic Vertical Gravity Gas and Liquid Separator